New gas liquefaction and other gas processing plants are being designed for ever-increasing production rates in order to realize the favorable economic benefits associated with larger plants. These larger plants have larger refrigeration duties with higher refrigerant circulation rates, and therefore larger refrigerant compressors are required. As gas processing plants become larger, the maximum achievable production rates may be limited by the maximum available compressor sizes.
When a single refrigerant compressor is used, these increased refrigerant flow rates require larger impellers with higher tip speeds, larger and thicker wall casings, and increased inlet velocities to the impellers. As the sizes of the compressor components are increased, the compressor will reach its fundamental aerodynamic limits, and this will fix the maximum possible compressor capacity. Many refrigeration systems utilize multiple refrigerant streams at different pressures, and these systems generally require compressors having multiple interstage suction inlets. The manufacturing and installation of these large, multistage compressors become significantly more difficult as compressor size increases.
A conventional multistage refrigerant compressor is illustrated schematically in FIG. 1. Refrigeration system 1 represents any type of refrigeration system in which multiple refrigerant streams are vaporized at different pressure levels to provide refrigeration in multiple temperature ranges. In this example, refrigeration system 1 utilizes four refrigerant streams that are vaporized in appropriate heat exchangers at four different pressures to provide refrigeration in four temperature ranges. Four vaporized refrigerant streams in lines 3, 5, 7, and 9, each at a different pressure, are withdrawn from system 1 and are introduced into the stages of multistage compressor 11 at the appropriate locations depending on the pressure of each stream.
The lowest pressure vaporized refrigerant in line 3 is introduced into the inlet of first stage 13, which may be designated as low pressure stage A. The low-intermediate pressure refrigerant stream in line 5 is introduced into second stage 15 of compressor 11, which may be designated as low-intermediate pressure stage B. The high-intermediate pressure refrigerant stream in line 7 is introduced into third stage 17 of compressor 11, which may be designated as high-intermediate pressure stage C. The high pressure refrigerant stream in line 9 is introduced into fourth stage 19 of compressor 11, which may be designated as high-pressure stage D. Each stage of the compressor may comprise one or more impellers and will compress an increasing mass flow of gas. Final compressed refrigerant gas returns via line 21 to refrigeration system 1.
The mass flow through low pressure stage A (first stage 13) is the mass flow entering in line 3; the mass flow in low-intermediate pressure stage B (second stage 15) is the sum of the mass flows entering in lines 3 and 5; the mass flow in high-intermediate pressure stage C (third stage 17) is the sum of the mass flows entering in lines 3, 5, and 7; and the mass flow in high pressure stage D (third stage 19) is the sum of the mass flows entering in lines 3, 5, 7, and 9.
When using single multiple-stage compressor 11 at a fixed driver speed, the total flow capability of the refrigeration system is limited by restrictions in the aerodynamic shape factors and flow factors which are used to design the compressor impellers. A speed reduction gear or a slower speed driver may eliminate these constraints in some cases. However, a speed reduction gear will add capital cost and result in mechanical power losses. Also, a speed reduction gear may complicate the mechanical torsional constraints of the compressor system and compromise the mechanical design of the system. The slower speed compressor stage in such a system will require larger casing sizes and larger impellers, which will add significantly to both the capital and installation costs. Thus the maximum size of single multiple-stage compressor 11 may be limited by any of these design factors.
Several alternative methods have been proposed in the art to compress large refrigerant flows in a multi-level refrigeration system. One solution is to use two identical half-size parallel compressors having a common inlet suction pressure source, common intermediate suction pressure sources, and a common outlet discharge pressure. The piping systems around the two parallel compressors must be meticulously designed and balanced so that both machines operate with the same flows through all stages of the compressors. Any flow imbalance between the two compressors will cause one of the units to reach surge (flow reversal) prematurely. Slight differences in manufacturing tolerances between the two machines, such as in the casings and impellers, will also contribute to flow imbalance.
Another alternative method to compress large refrigerant flows in a multi-level refrigeration system is disclosed in International Publication WO 01/44734 A2 and is illustrated in FIG. 2. In this alternative, the lowest pressure vaporized refrigerant in line 3 is introduced into the inlet of first stage 23, which may be designated as low pressure stage A, of first compressor 25. The high-intermediate pressure refrigerant stream in line 7 is introduced into second stage 27, which may be designated as high-intermediate pressure stage C, of first compressor 25. The low-intermediate pressure refrigerant stream in line 5 is introduced into first stage 29, which also is designated as low-intermediate pressure stage B, of second compressor 31. The high pressure refrigerant stream in line 9 is introduced into second stage 33, which may be designated as high pressure stage D, of compressor 11. Each stage of compressors 25 and 31 may comprise one or more impellers and will compress an increasing mass flow of gas. Final compressed refrigerant gas streams in lines 35 and 37 are combined and returned via line 39 to refrigeration system 1.
The mass flow through low pressure stage A (first stage 23) is the mass flow entering in line 3; the mass flow in high-intermediate pressure stage C (second stage 27) is the sum of the mass flows entering in lines 3 and 7; the mass flow in low-intermediate pressure stage B (first stage 29) is the mass flow entering in line 5, and the mass flow in high pressure stage D (third stage 33) is the sum of the mass flows entering in lines 5 and 9. This split compressor arrangement provides a method to eliminate the size and inlet velocity problems of single large compressor 11 (FIG. 1) without incurring the balancing problems of two identical half-size compressors discussed above.
Because gas liquefaction and other gas processing plants are being designed for ever-increasing production rates in order to realize the favorable economic benefits associated with larger plants, alternative methods are needed to eliminate the size and inlet velocity problems of single large compressors. Embodiments of the present invention, as described below and defined by the claims that follow, provide an alternative method for the design of refrigerant compressors for large gas liquefaction and processing plants.